Turbine driven pump



Mal-ch24, 1953 c. A. SCHELLENS TURBINE DRIVEN PUMP 7 Sheets-Sheet 1 Filed May 19, 1950 I-NVENTOR ATTORNEYS March 24, 1953 c. A. SCHELLENS 2,632,596

TURBINE DRIVEN PUMP Filed May 19. 1950 7 Sheets-Sheet 2 IN VENTOR 6. Said/ens ATTORNEYS March 24, 1953 c. A. SCHELLENS 2,632,595

TURBINE DRIVEN PUMP Filed May 19, 1950 7 Sheets-Sheet 5 1N VENTOR ATTORNEYS C. A. SCHELLENS TURBINE DRIVEN PUMP March 24, 1953 .7 Sheets-Sheet 4 Filed May 19, 1950 ATTORNEYS March 24, 1953 Filed May 19. 1950 C. A. SCHELLENS TURBINE DRIVEN PUMP 7 Sheets-Sheet 6 IN VENTOR 6. SCAc-M/eas Z6 BY ZZQZ M ATTORNEYS March 24, 1953 c. A. SCHELLENS 2,632,596

TURBINE DRIVEN PUMP Filed May 19, 1950 7 Sheets-Sheet '7 INVENTOR ScAe/kms ATTORNEYS Patented Mar. 24, 1953 1' UNITED STATES PATENT OFFICE TURBINE DRIVEN PUMP Christopher A.-Sche'llens, Tenants Harbor, Maine Application May 19, 1950, .SerialNo. 162,856

This application is a continuation-impart of Serial Number 637,337, filed December 27, 1945, now abandoned.

The efficiency of power plants by which is meant the ratio between output and input has been greatly improved in recent years by the adoption of higher and highersteam pressures. These improvements, however, tend to follow the law of diminishing returns, i. e. in order to make a substantial gain at the present time a large increase in steam pressure is required. Increasingsteam presures .throw a heavy burden on the boiler .feed pumps. In the first place the input to the fluid (assuming 1.00% pump efficiency) to increase its pressure form .an increasing percentage of the energy available in the steam cycle for producing work. Secondly, the high'pressures are a great handicap to efficient pump operation on account of the high windage and leakage losses. The result is that in increasing the pressure of thesteam cycle equilibrium is soon reached between the marginal thermodynamic gain-on the increased input and losses in the pump.

This invention relates to improvements in turbine driven pumps and especially to steam turbine driven centrifugal pumps suitable for pumping liquids. More particularly, the invention provides a pumping .unit of the mentioned general character wherein a plural-stage steam turbine-and a plurality of centrifugal pumps are combined and .coact to attain an efliciency greatly exceeding the efficiency of prior comparable pumping units of which I am aware.

.It is well'known in the art of steam turbine driven centrifugal pumps that where the rotative speedis high and where the liquid as supplied to the impeller eye is near its boilingv point, care must be taken. in the design of the impeller to avoid flashing or cavitation as defined below as the impeller eye. With a given volume and eye pressure, liquidcan be pumped at a higher temperature withoutdanger of cavitating as'the rotative speed of the impeller eye is decreased and its diameter increased.

A low rotative speed in a high, pressure-pump, however, requires large impeller diameter to maintain the rated'pressure rise. In .such impellers, the proportion of the power absorbed bythe surface friction of the walls to the useful power delivered in pumping rises rapidly as "the impeller diameter increases. Therefore, high pressure pumps ofthe type where the delivery volume isrelatively'small are, as at, present constructed, limited in efiiciency by the necessity or adopting a sufficiently low rotative speed to 3 Claims. (01. 230-5) avoid the danger of cavitation and, consequently, a large diameter impeller to maintain the required pressure rise.

Itis among the objects of my present invention to provide a turbine driven centrifugal pumping unit which is capable of maintaining a high suctioncapacity and which avoids the heretofore prevalent excessive vpower losses, due .to surface friction of large diameter impellers. I provide a plural-stage steam turbine whose first stage drives a high speed pump, and whose second stage drives a low speed ,pump which .delivers through the high speed pump. The impeller of the low speed pump may have small diameter with a large diameter eye, whereby its suction capacity will be high with low pressure rise and friction, and theimpeller of thehigh speed pump may'have a small. diameter with a small diameter eye, whereby it can receive the output of the low speed pump and deliver it with large pressure rise and low friction loss without danger of cavitation.

Another object is to provide a plural-wheel steam turbine having provision whereby jets of steam propel 1one wheel at high speed in one direction and propel another wheel at slower Speed in the opposite direction, thereby to drive respectively asecond-stage high speed pump and a first-stagelow speed pump, while at the same time securing important improvements in turbine efiiciency.

It is, moreover, my purpose and object generally to'improve thestructure, operation and efficiency of turbine driven centrifugal pumping units .and especially such .units of the steam turbine driven variety.

In the accompanying drawings:

.Figure l is .an elevation of .a steam turbine driven centrifugal pumping unit embodyingfeatures of my invention, the view being somewhat diagrammatic and exterior operating partsbeing omitted;

Figure 2 is a medial vertical cross-sectional view of substantially the left hand half of the unit of'Figurel, on a larger scale;

.Figure 3 is asimilar view of substantially the right hand'half of'theunit of Figure 1;

Figure 4 is ,a'fragmentary detail sectional view through a'plurality of thepropelling fluid nozzles and through a plurality of the buckets ofthe first andsecondstage turbine wheels;

FigureS is a cross-sectional view onlline 5-5 of Figure 2;

Figure 6 is an elevation of the emergency governor trip mechanism, looking towards the leftinFigure '5;

Figure 7 is a fragmentary sectional view on a larger scale showing the details of the mountin means for the pump impellers;

Figure 8 is an elevation, partly in section, and on a larger scale, showing one of the governors with its weights thrown outward;

Figure 9 is a sectional view through a slightly modified low pressure pump illustrating the specific construction thereof;

Figure 10 is an enlarged view showing the runner partly in elevation and partly in section of the pump shown in Figure 9; and

Figure 11 is a plan view of the intake side of the runner of the construction shown in Figures 9 and 10.

Referring to the drawings, and more particularly to Fig. l, a low pressure pump is indicated generally at I0, a high pressure pump at I2, and a plural-stage turbine at I l, the rotatable elements of all of which turn about a common axis. Liquid from any suitable source is conducted through conduit IE to the low pressure pump I and is delivered from pump I0 through conduit I8 to the high pressure pump I2 which delivers into conduit 20. Steam from any suitable source is conducted to the turbine I l through conduit 22, it passing from conduit 22 into the control valve unit indicated generally at 24, whence it passes to the turbine.

Referring now to Fig. 3, the low pressure pump I0 has an axial inlet 26 with which the liquid supply conduit I0 connects, and has the discharge outlet 28 delivering into conduit I8 which leads to the high pressure pump I2. Within the low pressure pump casing is rotatably mounted a relatively small diameter impeller 30 whose inlet 3 I is positioned and adapted to receive liquid from the pump inlet 26 and to conduct it to the relatively large diameter impeller eye 32. The impeller has a hub 34 secured to one end of shaft 30, which shaft is journalled in anti-friction bearings 38, 40 within the casing section 42. The impeller hub is equipped with a Wearing ring sleeve 40, and wearing rings 46, 38 surround sleeve 44 and inlet 3|, respectively, with close clearance, thereby restraining leakage from the impeller discharge. The usual annular chamber 50 surrounds the hub wearing sleeve 44, and the soft packing 45 seals chamber 50 against escape 'of liquid therefrom so that leakage into chamber 50 is conducted harmlessly through a conduit 52 (Fig. 1) leading from chamber 50 to the suction inlet 26, or its supply conduit I6, and the pressure in annular chamber 50 is maintained, through conduit 52, always the same as the pressure in the suction inlet 26. The usual diffuser 54 is arranged for coaction with the impeller 30.

A centrifugal emergency governor B is mounted on shaft 36 between its bearings 38, 40 for a purpose which later will appear.

The high pressure pump I2 is shown in detail in Fig. 2, it having an axial inlet 58 to which conduit I8 delivers, and a discharge outlet 60 delivering to conduit 20. Within the high pressure pump casing is rotatably mounted the relatively small diameter impeller 02 whose inlet $3 is positioned and adapted to receive liquid from the pump inlet 58 and to conduct it to the relatively small diameter impeller eye 64. Impeller 62 has a hub 66 secured to one end of a shaft 68 which shaft is journalled in anti-friction bearings I0, 12, and which is co-axial with the shaft 36 of the low pressure pump. The impeller hub 66 is equipped with a wearing ring sleeve 14, and wearing rings 16, I8 surround sleeve I4 and inlet 4 63, respectively, with close clearance, thereby restraining leakage from the impeller discharge. An annular chamber surrounds the hub wearing sleeve 14, and the soft packing 75 (Fig. 7) seals chamber 80 against escape of liquid therefrom so that leakage into chamber 80 is conducted through a conduit 82 (Fig. 1) leading from chamber 80 to the conduit 52 and thence to suction inlet 26, or its supply conduit I6, and the pressure in annular chamber 80 is maintained, through conduit 82, always the same as the pressure in the suction inlet 26. The customary diffuser 80 is located for coaction with impeller 62.

Shaft 08 has a centrifugal emergency governor 86 mounted thereon between the bearings I0, I2 of the shaft. Governor 86 may be of a known type including a flanged sleeve 88 which moves to actuate a rocker element 90 when centrifugal force causes weights 9| to fly outward (Fig. 8) in the event of an over-speeding of the high pressure rotary element of the turbine. Actuation of rocker element 90 may operate mechanism for effecting a slowing down or stopping of the overspeeding turbine element. As herein shown, the rocker element 90 operates mechanism indicated generally at 92 (Fig. 2) which effects stopping of the turbine by cutting off the steam supply, as more particularly described in my co-pending application, jointly with Arthur L. Sherman, Serial No. 512,584, filed December 2, 1943, now Patent No. 2,425,958, dated August 19, 1947.

The governor 50 on the low pressure motor shaft 36 similarly may include a sleeve 94 for actuating a rocker element 96 and, according to the invention, the rocker elements 90 and 96 of the two governors are connected together by a tie-link 98 (Fig. 1) so that any over-speeding of either the high pressure or low pressure rotary elements of the turbine will cause operation of mechanism, of which mechanism 92 is exemplary, for stopping or slowing down the turbine:

The turbine, indicated generally at I4, is located between the high and low pressure pumps. It comprises a first stage turbine wheel I00 and a second stage turbine wheel I02. Wheel I00 is fixed on the inner end of shaft 68 of the high pressure pump, and wheel I02 is fixed on the inner end of shaft 30 of the low pressure pump. The first or high speed stage wheel I00 has buckets I04 thereon, preferably of the impulse type, and the second or low speed stage wheel I02 has buckets I05 thereon closely adjacent to the buckets I 04 but in generally reversed arrangement. Since wheel I00 operates at a high speed, I prefer to construct it generally symmetrical with respect to a plane through the bucket centers in order to provide strength against rupture by centrifugal forces. The wheel I02 which operates at a lower speed may be of dished crosssection since the centrifugal forces inducing stresses in this wheel are not excessive.

A steam nozzle I08 is mounted within the turbine casing, being retained in position closely adjacent to the buckets I04 of wheel I00 by the retaining screws H0, at intervals around the easing. Nozzle I00 has the usual series of nozzle openings II2 therethrough for directing expansible rliuid into the buckets I04 thereby to cause high speed rotation of the first stage turbine wheel I00, and subsequently the fluid passes into propelling contact with the reversed buckets I06 of wheel I02, thereby to cause rotation of the second stage turbine wheel I02 in direction opposite to that of wheel I00. However, the fluid will have spent a considerable portion of its enorgy in propulsion of the first stage wheel I00, so that its velocity as it drives the second stage wheel m2 will be considerably reduced, and the fluid-will drive wheel I02 efficiently at a considerably reduced speed as compared with the speed of the first stage wheel Hill. Hence the high pressure pump impeller 62, to whose shaft 6 8 the first stage turbine wheel it!) is fixed, may be rotated at high speed in one direction efliciently, and the low pressure pump impeller 30, to whose shaft 3t the second stage turbine wheel I02 fixed, may be rotated at a considerably slower speed in the opposite direction while still maintaining its efiiciency, and the diameters of impellers 6.2 and ttlare proportional to secure the said rotative speeds.

The spent stream exhausts at the lower region of the turbine easing into exhaust outlet Ht which may be connected to a conduit Ht for leading it to any suitable-location.

In Fig. 4, there is indicated, somewhat diagrammatically in cross-section, a plurality of the nozzle openings i it of nozzle H1 8 in their relation to the buckets not, [0B of the turbine wheels we and H82, respectively.

'S-team may be supplied to the nozzle openings I t2 from any suitable source and in any suitable manner. As herein represented, the steam delivers into the turbine and to the nozzle openings a detailed description of the steam control valve mechanism herein generally indicated at 24. As described in the said application, the steam supply. to the turbine normally is automatically regulated and controlled by the valve mechanism H8 and its operating connections, to maintain a constant pump discharge pressure. The pressu-re'relief valve I120 in the pump discharge conduit 20 prevents the discharge pressure from rising-above a predetermined safe value and, since the speed of a pump impeller, assuming uninterrupted flow of liquid to its eye, depends upon the pressure against which it discharges, an overspeeding cannot occur in either pump if the inlet of the low pressure pump is open to a supply of liquid at a temperature sufficiently'abelow its boilingtemperature to avoid cavitation. (if there should be a failure of the supply of liquid, or if the temperature of the liquid should be excessively high, the supply of steam to the turbine would be shut off by the normal operation oft-he control valve mechanism 12d, as described inthe mentioned co-pen-ding application for patent. in such a case theemergency governor control would not operate to cause closing of the steam valve. It occasionally happens, however, that the normal operation of the steam control valve mechanism 2t, under unusually unfavora-b-le suction conditions, may permit a speed of either or both of the turbine wheels lllil, Hi2 slightly higher than may be desirable. Under such conditions, the emergency governors to and 8 6 respond to effect a closing of the steam control valve.

The specific structure of a high suction. liquid pump such as would be used in the present invention is shown in Figs. 9, 10 and 11. In these figures the drive shaft 235 corresponding to the drive shaft 3 6 is connected to the runner casting opposite the intake side with the runner having a hub 208 for this purpose.

lhe pump comprises the casing i210 having :a water inlet 226 to the pump bowl 212 around which there is the discharge chamber .2 It with an outlet :22'8 corresponding to the outlet :28 as shown in Figure 3 leading to the conduit I 8. The runner mechanism is shown at C.

The runner has an axial intake and a radical discharge from its periphery. It has a plurality of water passages a, b, c, d, e, f, and g terminating at the entrance or intake side in a plane perpendicular to the axis of rotation as shown in Fig. 11.

The intake ends of these passages are arranged around the central boss or crown 21-5, thewalls 2H5 between the passages at this point constituting cutting blades. The intake ends of the passages are thus segmental in section. The lower radial surfaces of the blades are the-driving surfaces at the entrance, the circumferential surfaces i2|1 and .2 t8 being the sidewalls of the passages at this .point.

However, the radical surfaces of the passages at the entrance spiral, gradually twist or change to lie in a plane at right angles to the axis of rotation and thus become the side walls of the passages toward and at the exit ends thereof. The circumferential surfaces which, at the entrance are the side walls of the passages gradually become the driving surfaces as the exit is approached. The passages gradually merge from a substantially segmental section at the intake ends thereof, into an approximately square or rather slightly oblong rectangular section and then gradually into a markedly oblong rectangular section at the exit. Hence, while the exit ends of the passages succeed one another at the periphery in the same plane, and while the intake mouths are also side by sid at the center, the passages overlie one another at intermedi ate points, as shown in Fig. 10.

This arrangement makes it possible to locate the intake ends quite close to the center and hence the velocity of the blades is comparatively low which, together with the smooth transition of driving from the radial to the circumferential surfaces before described, causes the water to enter at relatively low velocity without the formation of eddies. Although the water is discharged at high velocity and high pressure, the low entrance velocity and the smoothness of the drivingsurfaces prevent cavitation.

With reference to efficiency and the matter of leakage, hereinbefore mentioned, there is provided in this case an automatic, pressure operated packing for the runner which providesthe necessary clearance for the runner but holds such clearance at a minimum so that the loss through leakage is kept within allowable limits. Two of the packings are employed and, as they are preferably identical, only one as shown in Fig. 9 will be described. In its preferred form this packing comprises a ring 219 having inner flanges 229 and 22E extending from opposite faces of the ring. The lower flange'22l3 slidably fits in the sleeve 222 threaded into the bowl of the pump, and the ring can thus shift axially. Rotation of the ring is prevented by the pin 223. The external diameter of flange 22l is less than that of 21s, as stated, but the difference is so small as not to be readily apparent in Fig. 9. Otherwise, the floating ring would be held against the impeller, since the discharge pressure, acting on the surface external of the flange, would produce a greaterforce acting to the left than to the right, by virtue of'the greater area. If the two diameters were equal, the surfaces would just touch. I

The body of the ring lies in the pressure space of the bowl, and since the external diameter of the flange 22! is less than the external diameter of the flange 226, the area of the upper face of the body of the ring is greater than the area of the lower face. The upper face of the ring has an interrupted groove 224 leaving several ribs 225 having preferably the same depth as the flange 22l which tend to prevent cocking of the ring. The outer flange 226 of the ring has less depth than the flange 22L The operation is as follows. Assuming that the pump is at rest and the flange 221 is in contact with the runner, on starting up, the ring will be subject to unbalanced pressure which will move the ring away from the runner axially. As the ring moves a gap is opened between the flange 221 and the runner and the ratio of clearance as between the flange 22f and the runner to that between the flange 226 and the runner will increase, in consequence of which the pressure between the ring and the runner decreases. The same result could be obtained by drilling holes in the flange 228 instead of making it of less depth. Soon a point is reached where the force tending to push the ring away from the runner is balanced by the pressure on the back of the ring and the ring stays put" at a definite clearance from the runner. By proper proportioning, this clearance can be accurately determined in advance and I prefer to so proportion the parts as to maintain a clearance of from .001 to .002 of an inch. With this clearance, the leakage is not sufiicient to impair the efficiency of the pump in an amount which would render the pump unsuitable for the use intended, while at the same time rapid wear is avoided and the pump will remain in satisfactory operating condition for long periods. While I prefer to proportion flanges 220 and 22! so as to obtain a slight clearance between the ring 2E9 and the runner C as described above, I may so proportion these flanges as to permit the ring to touch the impeller with a light pressure between the surfaces in contact.

Having the shaft extend only partly into the runner from the face opposite the intake side thereof, as well as the overlying arrangement of the passages, makes it possible to arrange the inlet ends of the passages on a small radius.

The walls between passages at the entrance, are slightly chamfered.

As mentioned the construction above described and shown in Figs. 9, 10, and 11 constitutes a high suction liquid pump which will avoid cavitation as defined below.

In order to more clearly define the scope of applicants invention particularly with respect to the pumps which are involved, the following classification is well established in this art.

Centrifugal fluid pumps fall into three distinct classes, the structure of the pumps in each of which is governed by a different set of rules. These classes are:

Class 1.--Gas compressors in which the pressure rise is suflicient to produce a substantial decrease in the specific volume of the gas.

Class 2.(a Gas compressors (usually called blowers) which have a low pressure rise, insufificient to cause a substantial decrease in the specific volume of the gas.

(2)) Liquid pumps not subject to cavitation.

Class 3.Liquid pumps subject to cavitation.

The pumps described above and as called for in the following claims are those which fall in class 3.

Also the term cavitation as used in this application and set forth in the claims has the following limited meaning as taken from an article by G. F. Wislicensus in Mechanical Engineers Hand Book, Lionel S. Marks, fourth edition, pages 1897-1900, reading as follows:

The static pressure in a closed stream of fluid drops as the velocity of the flow is locally increased. The fluid velocities in a closed stream reach a definite upper limit as soon as the absolute pressure becomes equal to the vapor pressure of the fluid. When this limit is reached, the fluid vaporizes, forming vapor pockets in the stream which disturb the flow and, by their subsequent collapse, produce vibrations, noise, and destruction of the surrounding walls. This form of vaporization in a rapid stream of fluid is called cavitation.

In a centrifugal pump cavitation occurs if the inlet pressure is dropped below certain limits or if the capacity or speed of rotation is increased without a corresponding increase in inlet pressure. Cavitation in a centrifugal pump causes a drop in efficiency, vibration, noise, and a rapid deterioration of the impeller.

From the above definition it will be noted that cavitation as defined in this application and the following claims has reference only to liquids since only liquids have a vapor pressure.

Primary and essential requirements of the improved turbine-pump unit are that the turbine characteristics shall determine that a first stage turbine wheel will be propelled at high speed, and that a second stage turbine wheel will be propelled at a substantially lower speed, the second stage wheel driving a first stage pump impeller, and the first stage wheel driving a second stage pump impeller, with the first stage pump delivering through the second stage pump. Also important and essential is the absence of any mechanical drive connection between the respective pump impellers.

The lower speed turbine wheel is operated at a speed between one-fourth and two-thirds the speed of the higher speed turbine wheel or, the relative speeds of the wheels, namely the ratio of the high speed to the low speed wheel speed, lies within the range of 4 and 1 or 4/1 and 1 /1. The turbine wheels are propelled at high and relatively low speeds respectively and the pumps are designed to preserve substantially the inherent speed ratio of the turbine wheels. The low pressure first stage pump as shown for example in Figs. 9, 10 and 11 operates with maximum emciency at low speed and the other pump is designed for operation with maximum efiiciency at high speed. As a result both the turbine and pump operate with near maximum efficiency.

The shaft efficiency of a well designed pump stage is determined by the value wherein It isto be note'dthat an increase in K produces "an increase in shaft efliciency.

Since the capacity is usually predetermined,

two means for improving-theefiiciency are available, namely, increasing N and decreasing the pressure 'perstage, or, its equivalent, increasing thenumber of stages. When the latter expedient iscarried beyond a certain point the shaft becomes unduly'long, the consequent low critical speed resulting in vibrations, which soon destroy the fine clearances between impeller and wearing rings. lies in increasing N. This, however, can only The remaining avenue of improvement be done up to a certain point when cavitation as defined above occurs. Hence, the two speed construction. The speed of the first stage impeller being greatly reduced does not cause its efficiency to-suiier. When the second stage turbine drives the first stage pump, its proportion of the total pressure rise is small due to the turbine is the same for each, 1-! being the net positive suction head in feet, above vapor pressure at the eye. Therefore if. one impeller operates without cavitation the other will do so. As the above value increases it reaches a critical value K2 where both impellers cavitate. Through experimentation impellers of the type shown in Figs. 9., l0 and 11 have a high value of K2. Since H is usually predetermined it follows that N is limited by the highest known value of K2. The present invention requires a high value of K2.

In the turbine the greater part of the steam energy is absorbed by the first row buckets, and their peripheral speed is therefore made as high as wheel stresses will permit, taking into account overspeed margin necessary for feed pump servme.

In the construction shown when the peripheral speed of the second row of buckets is decreased a large gain in turbine efiiciency accrues which is an entirely unexpected result in the turbine art, also with this increase in efliciency the linear dimensions of the turbine decrease thereby permitting a more compact and light unit.

It is apparent that with a wheel speed ratio of second stage wheel speed to first stage wheel speed in the neighborhood of .5, the overall eificiency is very high while the size, determined largely by the small number of stages and turbine wheel diameters, is very small.

If the capacity is changed the R. P. M. is altered inversely as the square root thereof, and if the pressure range is increased the second stage pump is multistaged, the efficiency figures remaining about the same. The low pressure unit remains single stage for all ordinary conditions.

While I'have illustrated one particular steam control valve mechanism 25, as disclosed in the mentioned co-pending application for patent, it should be understood that any of various known types of steam control systems and structures may be employed with my novel pumping unitwithout departing from the scope of my invention as defined in'the appended claims. For example, various types of regulators responsive to pressure differences may be employed, or controls responsive to the speed of one or both of the pump shafts, or to other physical quantities applicable by those skilled in the art.

An impulse turbine as herein described, having nozzles I I2 providing substantially all of the expansion and discharging serially through the buckets of two turbine wheels which are propelled in opposite directions, has important advantages over conventional one-wheel turbines in which stationary reversing buckets are arranged between two rows of wheel buckets on a single turbine wheel. My improved turbine, by reason of the reversed rotation of the two wheels, eliminates the need for stationary reversing buckets and the consequent power losses incident to use of such buckets. Hence, the eificiency of my two wheel turbine inherently is high as compared with the mentioned conventional types. In the usual general applications of a turbine for driving a pump, the provision of two turbine wheels propelled in opposite directions would result in practical difficulties in transmitting the power to the usual single driven member of a pump. With my two wheel turbine, each driving a diiferent one of two coaxially arranged impellers of separate pumps, there is no difiiculty in transmitting power from the two turbine wheels to the pumps, and an extremely efiicient and practical combination is attained. This more especially is so when the second stage turbine wheel is connected for driving a low pressure pump impeller and the first stage turbine wheel is connected for driving a high pressure pump impeller, which is a preferable arrangement according to my invention.

Within practical limits of rotative speed and pump power, the highest turbine efliciency results when the speed of the second stage wheel is reduced considerably below the speed of the first stage wheel. Inasmuch as the total available energy of the expansible fluid is transformed into velocity in the nozzles, a larger proportion of this energy is spent in the first stage wheel, and the velocity of the steam as it reaches the second stage wheel is considerably less than its initial velocity at the first stage wheel. Hence, only a relatively slow speed of the second stage wheel is required in order to efficiently extract the remaining energy of the fluid. Consequently, a high over-all turbine efficiency is obtained, together with a high capacity for pumping hot water, due to the relatively slow speed or the second stage turbine wheel and the first stage pump which is directly driven thereby. Furthermore, it will be apparent that the power delivered to the second stage turbine wheel is considerably less than the power delivered to the first stage wheel, which has an important effect on the low pressure first stage pump efficiency, as presently will appear.

As previously explained herein, a high pump suction capacity demands low rotative speed of the pump impeller, and a high pump eirlciency demands high rotative speed of the pump impeller when the required total pressure rise is high. According to my present invention, I attain a high suction capacity of the low pressure first stage pump due to the relatively low rotative speed of its impeller. Since the power delivered for operating this pump is relatively small as compared with the power delivered for operating the high pressure second stage pump, the pressure rise of the first stage pump is but a fraction of the total pressure rise of the unit as a Whole, and a high efficiency of the lower pressure first stage pump is obtained because its relatively small pressure rise permits the use of a small diameter impeller, having a low friction loss at relatively low rotative speeds.

It will be obvious that, even though the temperature of the liquid entering the low pressure first stage pump impeller is near its boiling point, it will be greatly below the boiling point of the liquid when it is discharging from that pump due to the rise in pressure and corresponding rise in the boiling temperature of the liquid. Hence when the liquid delivers from the low pressure pump to the impeller of the high pressure pump, its temperature will be such, in relation to its boiling temperature, that the latter impeller may have a desired high rotative speed without any danger of cavitation. While the diameter of the high pressure pump impeller may be somewhat larger than that of the low pressure pump impeller, it nevertheless has a diameter which is regarded as small in the pump art, when regarded in relation to its pressure rise, and friction losses are reduced to a minimum with resulting high pum efiiciency. The high speed is particularly suitable in connection with the first stage turbine Wheel, attaining high efficiency of the combined unit.

It will be noted that the turbine described above is of the impulse type. The aforementioned favorable relation of the pump and turbine characteristics as well as the simple and practical turbine construction described applies only if the impulse type of bucket action is employed. The means necessary to secure impulse action are:

1. Buckets of such shape that the section area of the bucket passage at its discharge end is sufficient to pass the volume of steam entering the passage at a velocity which is equal to the spouting velocity taken relative to the bucket velocity, less an allowance for frictional retard tion through the bucket passage. This spouting velocity mentioned means the steam velocity resulting from a fully expanded nozzle discharging to the exhaust casing pressure.

2. A substantially fully expanded nozzle to the exhaust casing pressure. In actual practice the nozzle mouth may be reduced slightly to take care of thickness of bucket edges and Walls between adjacent nozzles.

3. Sufiicient leakage area from the entrance side of each bucket row to the exhaust casing to permit equalization to casing exhaust pressure at these points to provide for the effect of scale deposits in the bucket passages, underestimate of friction losses, etc. Otherwise the pressure might back up at the entrance side of the bucket causing a bucket reaction accompanied by an axial thrust on the wheel which would require additional structure.

When the turbine is of the pure impulse type and where the first stage impeller is proportioned to efiectively avoid cavitation, the favorable speed ratios called for in the specifications are structurally determined essentially by the selection of the proper ratios of impeller diameters. Thus cavitation limits the first stage pump to a certain R. P. M. which, taken with the selected speed ratio, assigns a value to the second stage pump R. P. M. Variation in the entrance and exhaust steam pressure having a relatively small effect on the spouting velocity of the nozzles. The diameter of the turbine wheels is selected to give the highest practical linear speed of the first row buckets from a strength standpoint. Therefore the ratio wheel speed/spouting velocity may be assumed to have an almost constant value. Taken together with the selected wheel speed ratio this determines the relative power generated by each of the two turbine wheels provided pure impulse action prevails in the buckets and this value must substantially equal the ratio of the squares of the linear speeds of the impeller peripheries. The ratio of the two impeller diameters taken with their assigned R. P. M. must produce this value.

It should be understood that in cases where the rated pump pressure is unusually high, either the low pressure pump or the high pressure pump, or both, may be divided into a suitable number of pump stages for attaining the required pressure rise, and that the herein described advantages applicable to the illustrated two-stage unit will accrue with multiple-pump stages.

I have described my invention in its preferred form, in connection with the illustrated preferred embodiment wherein the turbine is of the impulse type having two oppositely directed velocity stages in series. It is Well known, however, that in large turbine units where the highest possible steam economics are attainable and are commercially justified, the turbines take the form of the multi-pressure stage impulse type or the multi-stage reaction type. In the initial stages of such turbines, where the steam density is great, high wheel friction and interstage leakage losses occur, and it is advantageous to employ high rotative speeds and small wheel diameters, but in the final stages the steambecomes expanded and lower rotative speeds together with larger wheel diameters are employed to advantage.

In the larger units I may adopt the conventional impulse or reaction turbine drive in which the initial turbine stages are direct-coupled to the second pump and operate at a high rotative speed, and the final turbine stages are connected to the first pump, and operate it at a relatively low rotative speed. The two shafts may or may not be coaxial and their direction of rotation may or may not be opposed.

It will be seen that by this arrangement I secure all of the above described advantages applyin to the preferred form of my invention, except that due to the omission of the reversing turbine buckets. This latter feature ceases to be of interest since the multi-stage turbine structure afiords steam economics which in large installations equal and possibly exceed those possible in turbines constructed in accordance with my preferred form.

I claim:

1. An impulse type pumping unit for relative 1y large high pressure output of liquid while avoiding cavitation if and when the liquid within the unit may approach its boiling temperature, said unit comprising two shafts in end to end adjacent relation and free from mechanical driving connection with each other, a first stage turbine wheel having a single row of buckets thereon fixed on the inner end of one of said shafts, a second stage turbine wheel having a single row of buckets thereon fixed on the inner end of the other one of said shafts, the buckets of said second stage wheel being reversed with respect to the buckets of said first stage wheel, a series of nozzles arranged for directing expansible fluid into propelling relation to the buckets of said first stage wheel and thence directly into propelling relation to the buckets of said second stage wheel thereby to rotate the wheels in opposite directions, a first stage centrifugal pump having its impeller fixed to an outer portion of the shaft of said second stage wheel, a second stage centrifugal pump having its impeller fixed to an outer portion of the shaft of said first stage wheel, means connected between the two pumps for delivering the output of said first stage pump through the said second stage pump, said second stage pump impeller having a diameter at least as great as the diameter of the said first stage pump impeller and a relatively small diameter eye with respect to that of said first stage pump and the two said pump impellers having relative pumping characteristics for predetermining and maintaining the relative speed of said turbine wheels within the range of four to one to one and one-half to one, with said first stage turbine wheel operating at the higher speed.

2. In an impulse type turbine driven centrifugal pumping unit for pumping liquids while avoiding cavitation if and when the liquid within the unit may approach its boiling temperature, a pair of coaxial shafts in slightly spaced end to end relation and free from mechanical driving connection with each other, a pair of bucket-carrying members, each fixed on a different one of said shafts at the abutting ends of the shafts, each said member having a single annular series of buckets thereon, a turbine casing defining a fluid chamber, an annular series of nozzles adjacent the buckets of one of said members and in fluid-receiving relation to said fluid chamber, means for supplying an expansible fluid to said fluid chamber whence it passes to said nozzles, said nozzles being arranged and adapted to direct said fluid into propelling relation to the buckets of one of said members and thence directly into propelling relation to the buckets of the other said member, the latter having its buckets reversed with respect to the buckets of the first mentioned member, and being propelled by partially spent fluid at a speed substantially slower than and in direction opposite to the speed and direction of propulsion of the other member,

a first stage and a second stage centrifugal pump 7 each connected to a different one of said shafts,

runners in said pumps, one of said runners having a plurality of enclosed circumferential spaced water passages, said passages being of substantially rectilinear cross section of which two pposite sides are helicoidal with a varying helical pitch, and of which the other two sides are spiral with a varying spiral pitch, said runners being so proportioned as to differentiate the speeds of said bucket-carrying members, said runner in said second stage pump having a diameter at least as great as the diameter of the first stage pump runner and a relatively small diameter eye with respect to that of the first stage pump, and means connected to deliver the discharge of the slower speed pump through the other pump.

3. In an impulse type turbine driven pumping unit for pumping liquids while avoiding cavitation if and when the liquid within the unit may approach its boiling temperature, a first stage centrifugal pump having relatively high suction capacity at relatively low speeds, a second stage relatively high speed centrifugal pump receiving the output from the first stage pump, a first stage bucket-carrying turbine member driving said second stage pump, a second stage bucket-carrying turbine member driving said first stage pump, runners in said pumps, one of said runners having a plurality of enclosed circumferential spaced water passages, said runners being so proportioned as to differentiate the speeds of said bucket carrying members, the runner of said second stage pump and the runner of said first stage pump having substantially equal diameters and the second stage pump runner having a relatively small diameter eye with respect to that of the first stage pump, means for fluid propulsion of the first stage turbine member at relatively high speed, and for relatively low speed fluid propulsion of the second stage turbine member by the partially spent fluid issuing from the first stage turbine member directly into propelling relation to the second stage turbine member.

CHRISTOPHER A. SCHELLENS.

REFERENCES CITED The following references are of record in the file of this patent:

UNITED STATES PATENTS Number Name Date 706,187 Lindmark Aug. 5, 1902 717,877 Lindmark Jan. 6, 1903 760,035 Stumpf May 17, 1904 768,076 Rateau Aug. 23, 1904 855,809 'Rateau June 4, 1907 1,206,371 Pogue Nov. 28, 1916 1,664,488 Schellens Apr. 3, 1928 1,947,477 Lysholm Feb. 20, 1934 2,092,351 Huntzicker Sept. 7, .1937 2,234,733 Jendrassik Mar. 11, 1941 2,410,769 Baumann Nov. 5, 1946 FOREIGN PATENTS Number Country Date 560.584 Germany Oct. 4, 1932 

